Dispatchable storage combined cycle power plants

ABSTRACT

A dispatchable storage combined cycle power plant comprises a combustion turbine generator, a steam power system, a heat source other than the combustion turbine generator, and a thermal energy storage system. Heat from the heat source, from the thermal energy storage system, or from the heat source and the thermal energy storage system is used to generate steam in the steam power system. Heat from the combustion turbine may be used in series with or in parallel with the thermal energy storage system and/or the heat source to generate the steam, and additionally to super heat the steam.

CROSS REFERENCE TO RELATED APPLICATIONS

This application is a continuation of International Patent ApplicationNo. PCT/US2016/060608 titled “Dispatchable Storage Combined Cycle PowerPlants” and filed Nov. 4, 2016. PCT/US2016/060608 claims benefit ofpriority to U.S. Provisional Patent Application No. 62/251,338 titled“Dispatchable Solar Hybrid Power Plant Extended Operation” and filedNov. 5, 2015. Each of these applications is incorporated herein byreference in its entirety.

FIELD OF THE INVENTION

The invention relates generally to the operation of combined cycle powerplants with heat storage, including for example solar hybrid combinedcycle power plants having heat storage.

BACKGROUND

Electricity is a just-in-time′ commodity, necessitating that supply anddemand of power be balanced in order to maintain specified frequency andvoltage. The electric demand or load varies based on the needs ofconnected industrial, commercial, or residential customers for lighting,HVAC, electronics and appliances, pumps and motors, etc. Electricitydemand displays patterns that are influenced by the level ofmacro-economic activity, weather, customary working hours, time-of-day,as well as many other factors.

Electric generation must supply the demand, generally in the mosteconomic manner, given constraints on fuel cost/availability, powerplant operating/maintenance condition, availability of cooling for heatengines, and transmission costs or capacity constraints. Generatingunits are generally dispatched in merit order, based on the marginalcost of generation, with the most economical operated the most andsupplying the ‘base load.’ In addition to the marginal cost ofoperation, which is proportional to the cost of fuel and other variablecosts such as an overhaul accrual, generating units also have startupcost for fuel and wear-and-tear to bring them from a cold condition toan operating condition. Accordingly, some generating units may be placedin a ‘hot standby’ condition, if the standby operating costs are lessthan the startup costs, and the unit is required to meet anticipateddemand or reserve margins.

The marginal operating cost will determine when a generating unit isdispatched and how many megawatt-hours it will produce. In a regulatedintegrated utility responsible for generation, transmission anddistribution, the fixed costs for capital amortization are covered asreturn on invested capital mandated by utility commissions or other ratesetting entities. With the introduction of electric markets, independentpower producers (IPPs) must cover fixed costs and profit from thedifference between the price of electricity and the marginal cost togenerate power. This creates a tension between the efficiency of a newgenerating unit and the cost to build it, as novel efficiency measuresmust pay for themselves as well as a risk premium. The book value of apower plant is first cost less accumulated amortization offset byimprovements, but its market value may be for example determined as thenet present value of discounted cash flows, which depends on theinvestors' return requirements and forecasts of the cash flows. The cashflow forecasts are influenced by the economic dispatch order, which maychange based on fuel costs or if newer and more efficient units areconstructed.

Because sunlight is free, solar power plants have very low marginalcosts and are always at the front of the dispatch queue, and in somejurisdictions are mandated to be dispatched first. However solar thermalpower plants are relatively expensive to construct ($3 to $6 per watt,for example), and their low capacity factor (typically less than 25% ofnameplate), requires a high price for electricity to cover the fixedcosts and profit. As a matter of public policy, various countries andutility markets provide incentives to encourage construction of solarpower plants, using mechanisms such as:

-   -   renewable energy pricing policy such as feed-in tariffs (FITs),        standard offers, or power purchase agreements (PPAs) offering a        guaranteed payment per kilowatt-hour;    -   tax policy such as investment or production tax credits and        accelerated depreciation;    -   environmental policy such as carbon credits or taxes and        renewable portfolio standards.

The most important of these has been pricing policy, because it isrevenue that is most important in determining whether an investment in anew power plant will be profitable, and revenue certainty reduces thefinancial risk premium. Energy pricing policy has favored solar thermalpower plant designs that resemble base load plants with high efficiency,and has dis-favored load following capability. Feed in tariffs may alsofavor smaller power plants with higher capacity factors obtained bythermal energy storage. As renewable power has become competitive, theseincentives are being reduced or eliminated, and renewable power plantswill be expected to consider regulatory, market, commodity, andtechnology risks, similar to conventional power plants

Rankine Steam Cycles are commonly used to convert thermal to electricenergy. Raising the steam temperature tends to increase the powerconversion efficiency, permitting a smaller amount of heat (e.g., fossilor solar) to produce the same power. Regenerative feedwater heating alsoincreases efficiency, but at the expense of reduced power output for thesame steam flow. Raising the pressure tends to increase the specificwork (per unit of steam flow), permitting more power for the same sizepower block. Above pressures of about 50 to 75 bar, depending on thesuperheated steam temperature, the expanding steam may begin to condensewithin the turbine, potentially damaging the turbine by erosion.Accordingly at higher pressures, some form of reheat is required toavoid harmful condensation in the low pressure stage of the turbine.These and other factors determine the most economical steam conditionsand steam cycle for use in a thermal power plant.

Because solar fields are expensive, by conventional thinking it isdesirable to increase the thermal to electric efficiency of the powerblock by increasing the temperature and adding regenerative feedwaterheating. These steps result in the use of expensive, and sometimesexotic, materials, manufacturing, and construction measures. These stepstend to reduce the nameplate power output of the turbine-generator (perunit of steam flow), increase startup time, and reduce load followingcapability.

As more intermittent renewable generation is installed, load followingbecomes more important than base load power. Consequently there is anincreasing need for generating units with resiliency and flexibility tofollow load.

The “Duck Curve” illustrates the challenge of managing a green grid. Ascan be seen in FIG. 1, the California Independent System Operator(CAISO) Base Load is being reduced as solar power ramps up itsgeneration during the daytime. As the installed base of solar generationincreases over the years, the afternoon ‘hump’ disappears and becomes anincreasingly large depression causing two gigawatt-scale problems:

-   -   overgeneration risk occurs because thermal generation resources        must continue to operate to be available when the preferred        renewable resources become unavailable, as the sun sets for        example;    -   ramp need occurs as the evening load increase coincides with the        decreasing output from solar power.

Ironically, the conventional modern solar thermal power plant is notwell suited to this regime. Without thermal storage, such plantscontribute to the ‘Duck Curve’ depression. By adding thermal energystorage, conventional solar thermal plants can address over-generationby operating during the evening peak rather than during the afternoondepression. But their base load approach to power block design meansthese plants are not well suited to load following and cannot materiallyaddress the steep ramps.

SUMMARY

This specification discloses combined cycle power plants comprising athermal energy storage system that stores heat. The stored heat may beused in a bottoming cycle, for example to produce steam to drive a steamturbine in the bottoming cycle. The thermal energy storage system may becharged by any suitable source of heat, and may be discharged inmultiple passes, with a first pass discharging heat at a firsttemperature, and second and any subsequent passes discharging heat at asecond lower temperature and at subsequent still lower temperatures. Ifthe heat discharged from storage is used to produce steam to drive asteam turbine in the bottoming cycle, the steam turbine may be operatedin sliding pressure mode, with steam saturation temperature decreasingwith discharge temperature from the thermal energy storage system andsteam pressure decreasing with steam temperature. The steam producedfrom the thermal energy storage system may be superheated by heattransfer from the topping cycle exhaust. Discharging the thermal energystorage in this manner, combined with sliding pressure operation of thesteam turbine, allows more complete use of the heat stored in thethermal energy storage system.

As non-limiting examples of such combined cycle power plants, thisspecification discloses solar hybrid power plants that utilize bothsolar energy and fossil fuels to generate electricity, and methods ofoperating those power plants, that may increase the solar energyfraction compared to conventional hybrid solar-fossil power plants toreduce the overall heat rate and carbon emissions, integrate solarthermal energy collection and storage with fast-start combustionturbines to provide dispatchable solar power with load followingcapability, and reduce the overall cost of solar thermal power. This maybe achieved by synergistic arrangements of technologies, which mayinclude the following.

-   -   Integration of the solar derived steam into the combined cycle        using the solar thermal energy primarily or exclusively for        latent heat transfer (evaporation), while reserving the heat        from the turbine exhaust gas primarily or exclusively for        sensible heat transfer (feedwater heating and steam        superheating). Here “primarily” means ≥50%, ≥60%, ≥70%, ≥80%,        ≥90%, ≥95%, or ≥99%. Eliminating or reducing extraction steam        and feedwater heaters reduces first cost and enhances the load        following capability of the power plant.    -   Integration of the solar derived steam into the combined cycle        using the solar thermal energy primarily or exclusively to        produce a first stream of saturated or slightly superheated        steam, parallel generation of a second stream of saturated or        slightly superheated steam using heat primarily or exclusively        from the turbine exhaust gas, and production of superheated        steam from the combined first and second streams of saturated or        slightly superheated steam using heat primarily or exclusively        from the turbine exhaust gas stream.    -   Use of low vapor pressure heat transfer fluids (HTFs), which        permit thermal energy to be stored in low cost atmospheric        storage tanks. Suitable HTFs may include paraffinic heat        transfer fluids such as Duratherm 630™, for example, which are        less hazardous and less expensive than the HTFs used in typical        parabolic trough technology. The latter HTFs are also typically        unsuitable for energy storage because their high vapor pressure        would require very large and expensive pressure vessels.    -   Use of a low pressure (e.g., ≤about 75 bar, ≤about 70 bar,        ≤about 65 bar, ≤about 60 bar, ≤about 55 bar, ≤about 50 bar,        ≤about 45 bar or ≤about 40 bar) Rankine steam cycle with live        steam conditions of, for example, about 70 bar/470° C. or about        60 bar/550° C., or about 45 bar/470° C., for which the boiling        point is suitable for the low vapor pressure HTF, and for which        the combustion turbine-generator (CTG) exhaust gases are readily        able to provide the necessary superheat. These steam conditions        may also eliminate the need for reheat, because the turbine        exhaust moisture content may remain within acceptable limits,        and may provide further cost savings by permitting the use of a        single casing steam turbine, rather than separate high and low        pressure turbines.

These arrangements may overcome several technical limitations withconventional approaches to solar thermal power plants, such as forexample the following.

-   -   The maximum capacity of a conventional stand-alone solar thermal        collection system is generally greater than the capacity of the        steam turbine generator. Coupling a conventional thermal energy        storage system will generally improve the capacity factor of the        solar thermal system, but may not improve the economic utility        because of the expense of the storage system, and the reduced        steam temperature and cycle efficiency when operating from        stored energy.    -   In a typical conventional solar-fossil hybrid combined cycle        power plant designed to also operate without solar input, only        about 5% of the gross electric power can be derived from solar        thermal energy because flow and temperatures become sub-optimal        when the solar steam is excessive. In contrast, solar hybrid        power plants described in this specification may permit a 25% or        larger fraction of electric power to be derived from the solar        energy.    -   In a typical solar thermal plant employing parabolic trough        technology, synthetic HTFs are required in order to deliver live        superheated steam temperature of about 370° C. to achieve the        needed Rankine cycle efficiency. In solar hybrid power plants        described in this specification, more superheat may be obtained        via the CTG exhaust gases, enabling higher efficiency. Further,        the expense and environmental hazards of fluids such as        Therminol® may be avoided.    -   Molten salt is typically used for thermal energy storage in a        conventional solar thermal power plant because the synthetic HTF        is too expensive to use for storage, and the vapor pressure of        the synthetic HTF would require excessively large and expensive        thick-walled pressure vessels. Molten salts require special and        expensive materials, and expensive and energy consuming heat        tracing to avoid freezing and prevent obstructions in the molten        salt piping, instrumentation and valves. In addition, the salts        require an additional heat exchanger which adds cost and reduces        efficiency. By using low cost, non-toxic HIT with low vapor        pressure, solar hybrid power plants described in this        specification may avoid these difficulties.

These and other embodiments, features and advantages of the presentinvention will become more apparent to those skilled in the art whentaken with reference to the following more detailed description of theinvention in conjunction with the accompanying drawings that are firstbriefly described.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows plots of the California Independent System Operator BaseLoad as a function of time of day for Mar. 31, 2012 and 2013 (actualbase load), and for Mar. 31, 2014-2020 (projected base load),colloquially known as ‘the Duck Curve.’

FIG. 2 schematically illustrates an example solar hybrid power plant.

FIG. 3 schematically illustrates another example solar hybrid powerplant.

FIG. 4 plots temperatures for steam and heat transfer fluid duringdischarge of heat from a thermal energy storage system for nameplate andpart load operation of an example solar hybrid power plant.

FIG. 5 schematically illustrates an example configuration for a thermalenergy storage system.

FIG. 6 schematically illustrates a dispatchable storage combined cyclepower plant similar to the solar hybrid power plant of FIG. 3, with thesolar thermal system replaced by an electric heating system.

DETAILED DESCRIPTION

The following detailed description should be read with reference to thedrawings, in which identical reference numbers refer to like elementsthroughout the different figures. The drawings, which are notnecessarily to scale, depict selective embodiments and are not intendedto limit the scope of the invention. The detailed descriptionillustrates by way of example, not by way of limitation, the principlesof the invention. This description will clearly enable one skilled inthe art to make and use the invention, and describes severalembodiments, adaptations, variations, alternatives and uses of theinvention, including what is presently believed to be the best mode ofcarrying out the invention. As used in this specification and theappended claims, the singular forms “a,” “an,” and “the” include pluralreferents unless the context clearly indicates otherwise.

As noted in the summary, this specification discloses combined cyclepower plants comprising a thermal energy storage system that stores heatfor use in the power plant's bottoming cycle. Such combined cycle powerplants may be referred to in this specification as Dispatchable StorageCombined Cycle (DSCC) power plants. The bottoming cycle in a DSCC powerplant may be, for example, a Rankine cycle employing a steam turbine.

In the DSCC power plants described in this specification, the thermalenergy storage system may store heat in any suitable storage medium atany suitable temperature or range of temperatures. For example, thethermal energy storage system may store heat in a low vapor pressureheat transfer fluid (e.g., at a pressure of about one atmosphere) suchas, for example, an (e.g., organic) oil or an inorganic molten salt, orin a high vapor pressure heat transfer fluid, or in any other suitableliquid. The thermal energy storage system may store heat in a solid suchas, for example, graphite, metals, concrete (e.g., HEATCRETE provided byHeidelberg Cement AG and used in Energy Nest concrete storage modules),or aggregate. The thermal energy system may store heat in a gas or vaporsuch as, for example, steam. The thermal energy storage system may storeheat as sensible heat, as latent heat (i.e., via a phase change), or asboth sensible heat and latent heat.

The thermal energy storage system may be charged by any suitable sourceof heat. For example, solar energy may be collected as heat and storedin the thermal energy storage system, electric heaters (e.g., resistiveor inductive electric heaters) may be used to provide heating forstorage, or a fuel (e.g., a fossil fuel) may be combusted to produce theheat. If electric heaters are used, they may be powered with electricitygenerated with photovoltaic solar cells, with a solar thermal electricpower plant, with a wind turbine, with a hydroelectric power plant,using nuclear power, by combusting a fossil fuel, by any suitablecombination thereof, or by any other suitable method. The thermal energystorage system may be heated directly (e.g., by embedded electricheaters), or indirectly by a heat transfer fluid that transfers heatfrom the heat source to storage. In the latter case, suitable heattransfer fluids may include, for example, molten salts, steam, and oils,including high vapor pressures oils typically used by parabolic troughconcentrating solar power systems.

The thermal energy storage system may be discharged in multiple passes,with a first pass discharging heat at a first temperature, and secondand any subsequent passes discharging heat at a second lower temperatureand at subsequent still lower temperatures. The heat discharged fromstorage may be used, for example, to produce steam, which is thensuperheated by heat transfer from the topping cycle, to drive a steamturbine in the bottoming cycle, with each pass producing steam atsuccessively lower pressures. In such cases the steam turbine may beoperated in sliding pressure mode, with saturated steam temperaturedecreasing with discharge temperature from the thermal energy storagesystem and steam pressure decreasing with steam temperature. Althoughthe saturation temperature of the steam changes as the pressure slides,heat transfer from the topping cycle, and desuperheating sprays asneeded, maintain the temperature of the superheated steam to the steamturbine.

As examples of DSCC power plants and their operation, this specificationdescribes several dispatchable solar hybrid combined cycle power plantscoupled with energy storage. These examples are meant to beillustrative, not limiting. That is, a DSCC power plant need not utilizesolar energy. Instead, as noted above a DSCC power plant thermal energystorage system may be charged with heat from any suitable source.

When this specification refers to dispatchable solar hybrid combinedcycle power plants, the hybrid aspect means that power may be derivedfrom both solar and fossil energy sources. The combined cycle aspectmeans that power may be generated by both a Brayton topping cycle and aRankine bottoming cycle. The dispatchable aspect means that thermalenergy may be stored in order to produce power when the sun is notavailable.

FIG. 2 schematically illustrates an example solar hybrid power plant 10comprising a combustion turbine generator 100, a heat recovery system200, a solar thermal system 300, a steam power system 400, and a thermalenergy storage system 500. As described in more detail below, heat fromthe combustion turbine generator exhaust gas collected in heat recoverysystem 200 and heat collected in solar thermal system 300 may be usedseparately or in combination to drive steam power system 400. Heat fromsolar thermal system 300, and optionally heat from other sources, may bestored in thermal energy storage system 500 for later dispatch to steampower system 400. Operation of solar hybrid power plant 10 is describedbelow with reference to particular example temperatures, pressures, andflow rates. Any other suitable values for these parameters may also beused.

Referring again to FIG. 2, combustion turbine generator 100 may be aconventional combustion turbine generator operating on an open Braytoncycle in which atmospheric air is compressed and mixed with fossil fuelwhich is burned to produce hot gas that expands across a turbine. Thecompressor, turbine, and generator may be aligned along a rotatingshaft, optionally with an intermediate gearbox to match the shaftspeeds. Auxiliaries to the combustion turbine generator may include, forexample, inlet air filtering and cooling systems, lubrication systems,control and condition monitoring systems, and ventilation and firesuppression equipment. The combustion turbine generator may be, forexample, a Siemens SGT6-8000H, which is a modern 60 Hertz system with agross rating of 274 Megawatts at 40% efficiency. Any other suitablecombustion turbine generator may be used instead. Combustion turbine 100may exhaust, for example, about 604 Kg/s of exhaust gas at about 617° C.

Exhaust gas from combustion turbine generator 100 is directed into heatrecovery system 200, which extracts heat from the exhaust gas for use bythe steam power system 400. Heat recovery system 200 includes heatrecovery units (heat exchangers) disposed in the duct work connectingthe combustion turbine generator with the chimney/stack of the powerplant. The heat recovery units typically transfer heat from the exhaustgas to a fluid flowing through a coil.

Exhaust gas from the combustion turbine generator first flows throughhigh temperature heat recovery unit (HTHRU) 210, which functions as asuperheater to increase the temperature of steam generated in boiler 470(further described below) while cooling the exhaust gas to, for example,about 405° C. The exhaust gas may then optionally flow through exhaustgas treatment systems such as a selective catalytic reduction (SCR) 220system to reduce emissions of nitrogen oxides or a carbon monoxidecatalyst. After the optional SCR 220, the exhaust gas flows throughmedium temperature heat recovery unit (MTHRU) 230, which functions as aneconomizer to warm feedwater for boiler 470 while further cooling theexhaust gas to, for example about 160° C. The exhaust gas then flowsthrough low temperature heat recovery unit (LTHRU) 240, which capturesheat used to preheat the boiler feedwater prior to passage of the boilerfeedwater through MTHRU 230 and further cools the exhaust gas to, forexample, about 60° C. The exhaust gas is then discharged up thestack/chimney.

Solar thermal system 300 concentrates sunlight onto a heat collectorthrough which a heat transfer fluid flows to receive energy in the formof sensible or latent heat. Solar thermal system 300 may be or comprise,for example, a linear focus system using parabolic trough or compactlinear Fresnel reflector technology, or may be a dish or tower typeconcentrating solar thermal energy collector. Solar thermal system 300may employ, for example, Solar Reserve Inc. molten salt tower solarenergy collectors, Bright Source Energy, Inc. direct steam tower solarenergy collectors, AREVA Solar Inc. CLFR solar steam generators, orconventional parabolic trough heat collector elements. Any suitablesolar thermal energy collectors may be used in solar thermal system 300.Intermediate heat transfer fluid storage tanks and pumps (not shown) maybe arranged within solar thermal system 300 to facilitate operations andmaintenance.

Thermal energy storage system 500 stores thermal energy as sensible heatin a low vapor pressure liquid heat transfer fluid. As used herein theterm “low vapor pressure liquid heat transfer fluid” refers to a heattransfer fluid having a vapor pressure less than about 0.1 atmosphere,or less than about 0.5 atmosphere, or less than about 1.0 atmosphere, ata temperature of about 300° C., or about 330° C., or about 350° C. Hotheat transfer fluid stored in thermal energy storage system 500 may beused when needed to provide heat to steam power system 400 (e.g., toboiler 470).

The use of low vapor pressure heat transfer fluid in thermal energystorage system 500 permits the hot liquid to be stored in thin-wallinsulated tanks at about atmospheric pressure. In contrast, conventionalparabolic trough based solar thermal systems typically use fluids suchas Therminol®, which are not suitable for thermal energy storage becausetheir vapor pressure would require the use of pressure vessels, whichare impractical for the large volumes that may be desired.

As noted above in the summary section, one heat transfer fluid suitablefor use in thermal energy storage system 500 is the paraffinic heattransfer fluid Duratherm 630™, which is both less toxic and lower costthan the commonly used solar thermal fluids such as Therminol®.Duratherm 630™ has a vapor pressure of about 0.1 atmospheres at about316° C. Caloria HT-43 (a petroleum distillate) is another suitable heattransfer fluid. Other heat transfer fluids with low vapor pressure mayalso be used, including other paraffinic heat transfer fluids and moltensalts, subject to technical and economic considerations related to heattransfer fluid cost, the cost of compatible piping, tanks, valves andpumps, environmental and fire hazards, the freezing point and the costof heat tracing for low temperature operation or extended non-operatingperiods, the stability of the heat transfer fluid at high temperature,and the quantity of fluid required, based on the density and specificheat.

Sensible heat may be added directly to the low vapor pressure heattransfer fluid in thermal energy storage system 500 by solar thermalsystem 300 by pumping cold heat transfer fluid through solar thermalsystem 300. A control system may regulate the flow rate of the heattransfer fluid through the solar thermal system to achieve a desiredtemperature.

Alternatively, or in addition, sensible heat may be added indirectly tothe low vapor pressure heat transfer fluid in thermal energy storagesystem 500 from solar thermal system 300 by pumping cold heat transferfluid through a heat exchanger to receive heat from a higher temperaturefluid circulated through solar thermal system 300. This may involve, forexample, sensible heat transfer from a high vapor pressure heat transferfluid (such as Therminol®) heated to about 400° C. in a conventionalparabolic trough solar energy collector, or sensible heat transfer froma molten salt in the case of tower or novel linear technologies.Alternatively, this could involve latent heat transfer from steamcondensation. For example, solar steam condensing at about 132 bar wouldheat the low-vapor pressure heat transfer fluid to about 332° C., whichis the maximum operating temperature of Duratherm 630®.

Such indirect heating may reduce technology risk by using provencommercial technology and may also be a cost effective means ofupgrading an existing solar thermal power plant to a dispatchable hybridplant.

Sensible heat may also be added to the heat transfer fluid in thermalenergy storage system 500 by non-solar sources. Non-solar heating mayinvolve pumping the low vapor pressure heat transfer fluid through aheater or heat exchanger, for which the source of energy may be orinclude, for example, hot combustion exhaust gases from combustionturbine generator 100, or hot combustion gases from another gas turbineco-located with solar hybrid power plant 10. For example, a secondcombustion turbine generator could be provided for peaking service.Instead of utilizing a heat recovery steam generator, the exhaust gasfrom the second combustion turbine generator could flow across heatexchangers to transfer heat from the hot exhaust gas to a low vaporpressure heat transfer fluid which may then be stored in tanks. Thisarrangement could supplement the energy stored by a solar collectionsystem, or could be employed instead of a solar collection system. Thelatter case is similar to a two-on-one combined cycle plant, where twocombustion turbine generators and heat recovery steam generators feed asingle steam power cycle, but provides for time shifting of the powergeneration in the bottoming cycle. The low vapor pressure heat transferfluid could be heated directly in the exhaust stream or indirectly by,for example, condensing steam produced by a heat recovery steamgenerator.

Alternatively, non-solar heating may use for example a fossil fuelburner, process heat, or electric heating. In the latter case theelectric heating of the low vapor pressure heat transfer fluid may storeelectricity generated at remote locations at inopportune or uneconomictimes, with discharge at more favorable times. The electric heating maybe produced, for example, from solar PV or wind generation, includingover-generation as shown in FIG. 1.

Hot and cold low vapor pressure heat transfer fluid in thermal energystorage system 500 may be stored for example in dedicated hot and coldtanks (as illustrated), in one or more thermocline tanks in which hotheat transfer fluid is stored above cold heat transfer fluid in the sametank, or in tanks that may be used alternatively for hot or coldstorage. The latter two approaches may be suitable if the sensibleheating and cooling of the tank and associated piping is inconsequentialcompared to the economic savings achieved by reducing the number oftanks required. Filling, draining, and ullage of tanks and tank farmsmay follow conventional practice.

The hot and cold low vapor pressure heat transfer fluid in thermalenergy storage system 500 is typically stored at about atmosphericpressure. Thermal energy storage system 500 may operate between about330° C. (hot heat transfer fluid) and about 250° C. (cold heat transferfluid). Any other suitable temperature range may also be used.

Referring again to the example of FIG. 2, when heat is available fromsolar thermal system 300 a pump 540 may draw cold heat transfer fluid ata temperature of, for example, about 265° C. from low temperaturestorage tank 530, pump it through solar thermal system 300 for heatingand then to high temperature storage tank 510, where it is stored at atemperature of, for example, about 330° C. Pump 520 may pump hot heattransfer fluid from high temperature storage tank 510 through boiler 470to heat boiler feedwater to generate steam, and then back to lowtemperature storage tank 530. Alternatively, or in addition, heattransfer fluid may be pumped from boiler 470, through solar thermalsystem 300 for heating, and then back to the boiler to generate steam.That is, one or both of the heat transfer fluid storage tanks mayoptionally be bypassed. Hot heat transfer fluid may be supplied toboiler 470 in various proportions as desired from hot storage tank 510or directly from solar thermal system 300.

In steam power system 400, superheated steam (generated as describedbelow) is delivered to steam turbine generator (STG) 410, for example atdesign conditions of about 70 bar/470° C. and a flow rate of about 250kg/s. This design condition allows use of a compact single-casingturbine without reheat, because the expanded steam will not haveexcessive moisture (liquid content) in the last stage of the turbine.Reheating of the steam may be avoided to save cost, and elimination orreduction of extraction steam for feedwater heating enhances loadfollowing and permits all or more of the steam to perform useful work.

After expanding through the turbine, steam is condensed to liquid water,typically using an air-cooled condenser 420 at an annual averagecondensing temperature of, for example, about 40° C., which correspondsto an exhaust pressure of about 0.085 bar. Under these conditions, asteam flow of about 250 kg/s may generate approximately 225 Megawatts ofelectric power.

Condensing temperature varies depending on the cooling capacity, withlower condensing temperature permitting higher net power generation.Water cooling, if available, would increase the net power generation byreducing the condensing pressure, and eliminate or reduce power consumedby the fans of the air cooled condenser.

Heat derived from the combustion turbine generator exhaust gases usinglow temperature heat recovery unit 240 described above is used to raisethe condensate temperature from about 40° C. to about 105° C., forexample. At an example water (steam) flow rate of about 250 kg/s thismay require, for example, about 65 MW of heating. The condensate may beeither directly heated by circulating it through low temperature heatrecovery unit 240 or indirectly heated.

In the illustrated example, a condensate pump 430 transfers thecondensed water from condenser 420 through an optional condensatepolishing unit 440 and into a feedwater heater 450. The condensate isthen indirectly heated with a low temperature heat transfer loop 600that circulates a heat transfer fluid through low temperature heatrecovery unit 240 to collect heat and delivers the collected heat tofeedwater heater 450. Feedwater heater 450 may be for example a closedfeedwater heater. Alternatively, feedwater heater 450 may be an opendeaerating feedwater heater and the heat transfer fluid in loop 600 maybe used to produce low pressure pegging steam delivered to the feedwaterheater. Pegging or heating steam may in addition or alternatively bedrawn from any other suitable location in steam power system 400.

Indirect feedwater heating as just described may provide operationalflexibility to overcome transient mismatches between the heat availablein the combustion turbine generator exhaust gas and the heat requiredfor feedwater heating, particularly during startup and load following.For example, optional intermediate storage tanks in low temperature heattransfer loop 600 may hold heat transfer fluid to manage such transientmismatches between the availability of heat in the exhaust gas and theneed for feedwater heating.

Referring once more to the example of FIG. 2, boiler feedwater pump 460supplies water from feedwater heater 450 to medium temperature heatrecovery unit 230 at a pressure of about 75 bar, for example. Mediumtemperature heat recovery unit 230 raises the temperature of thefeedwater from about 105° C. to about 255° C., for example, by heattransfer from the combustion turbine generator exhaust gas.

The heated water then enters boiler 470, where it evaporates to producesaturated steam at a temperature of about 288° C. and a pressure ofabout 70 bar, for example. At a water (steam) flow rate of about 250kg/s this may require, for example, about 425 MW of heating providedfrom solar thermal system 300, from thermal energy storage system 500,or from a combination of solar thermal system 300 and thermal energystorage system 500. This would require about 2350 kg/s of heat transferfluid if Duratherm 630™ were used to supply heat to boiler 470 at aninput heat transfer fluid temperature of about 330° C. and an outputheat transfer fluid temperature of about 265° C.

Boiler 470 may utilize for example a once-through arrangement tofacilitate rapid startup and load following, or a recirculatingdrum-type boiler may be used. Boiler 470 may optionally superheat thesteam slightly, to about 290° C. for example.

The steam generated in boiler 470 flows to high temperature heatrecovery unit 210, where the steam is further superheated to atemperature of, for example, about 470° C. This superheated steam isexpanded across steam turbine generator 410 to generate power, and thencondensed as described above.

As just described, solar hybrid power plant 10 may operate as a combinedcycle power plant generating electric power with combustion turbinegenerator 100 and additional electric power with steam turbine generator410. The combustion turbine generator power output is determinedprimarily by the fuel flow, and may be rated for example at about 275MW. The steam turbine generator power output is determined primarily bythe steam flow.

When there is energy available from thermal energy storage system 500and/or solar thermal system 300 to heat boiler 470, the combustionturbine generator exhaust gas may be used primarily for single phaseheat transfer to warm water before it enters boiler 470, and tosuperheat steam produced in boiler 470, as described above. Under thesecircumstances steam flow to boiler 470 may be about 250 kg/s atconditions of about 70 bar/470° C., for example, and the steam turbinegenerator electric power output may be about 225 MW. With a 275 MWcombustion turbine generator the overall plant may then produce about500 MW of electric power, for example.

Power plant 10 may also be operated as a combined cycle power plantwithout using the thermal energy storage system, by bypassing boiler 470and for example using medium temperature heat recovery unit 230 and hightemperature heat recovery unit 210 in series as a once through steamgenerator. In such cases the steam flow to the steam turbine generatorwill be reduced, for example from about 225 kg/s with boiler 470 toabout 100 kg/s without boiler 470. With the lower flow rate but the samesteam turbine throttle temperature of about 470° C., the steam turbineinlet pressure is reduced from about 70 bar to about 30 bar and theelectric power output is reduced from about 225 MW to about 80 MW, sowith a 275 MW combustion turbine generator the overall plant producesabout 355 MW of electric power, for example.

Power plant 10 may also be operated as a combined cycle power plant atintermediate steam turbine generator operating conditions, with somesteam produced by boiler 470.

Power plant 10 may also be operated as a simple cycle power plant, bynot flowing water into heat recovery system 200, which is acceptable forsome once through boiler designs, or by bypassing the combustion turbinegenerator exhaust gas around heat recovery system 200. In this case, thepower plant produces electric power with the combustion turbinegenerator but not with the steam turbine generator. The overall powermay then be, for example, about 275 MW.

It is important to note that fuel consumption is unchanged in theseveral operating modes just described, so the cost of power ($/MW-hour)and the heat rate (BTU/kw-hr) increase as the power output decreases.

In another embodiment of a dispatchable solar hybrid combined cyclepower plant, shown in FIG. 3 and referred to as power plant 20, thefeedwater from pump 460 is split into two streams, one of which flows toMTHRU 230 and the other of which flows to Boiler 470. Both streams areboiled, one using heat from the exhaust gas and the other using heatstored in the low vapor pressure heat transfer fluid. In thisembodiment, MTHRU 230 and boiler 470 are arranged in parallel in thesteam circuit, and MTHRU 230 produces steam at about the sametemperature and pressure as that produced in boiler 470, that is,saturated or slightly superheated steam. Boiler 470 may produce steam atfor example about 3 to about 6 times the rate at which steam is producedin parallel in MTHRU 230. Steam from Boiler 470 is mixed with steam fromMTHRU 230 and enters HTHRU 210, where the mixed steam is superheatedbefore entering Steam Turbine 410.

Tables 1A-1G provide example flow rates, pressures, temperatures, andother parameter values for operation of power plant 20 described above.Any other suitable values for these parameters may also be used.

In this embodiment heating duty is displaced from MTHRU 230 to boiler470. By shifting the heating duty in this manner, more of the energystored in the low vapor pressure heat transfer fluid can be extractedbecause the heat transfer fluid can be cooled to a lower temperature,potentially reducing the volume of heat transfer fluid to be stored intanks. In addition, shifting heating duty to boiler 470 can reduce theaverage temperature of heat transfer fluid circulating in the solarfield and consequently reduce heat losses.

This embodiment may simplify the design and operation of MTHRU 230,because it always includes an evaporating section, whether or not energyis available from the energy storage system.

In power plant 10 (FIG. 2) and power plant 20 (FIG. 3), without steamproduction at normal operating temperature and pressure from the heattransfer fluid (e.g., after all of the hot heat transfer fluid instorage is consumed), the steam turbine inlet pressure may be reducedduring sliding pressure operation to maintain approximately constantvolumetric flow through the turbine. At the reduced turbine throttlepressure, the saturation temperature of steam is also reduced, forexample to about 234° C. at about 30 bar.

Alternatively, or in addition, additional thermal energy may beextracted from the heat transfer fluid during sliding pressureoperation. After heat has been extracted from the heat transfer fluid toevaporate steam during full pressure operation, the temperature of theheat transfer fluid may still be sufficient to boil additional steam atthe reduced operating pressure during part load operation. This may bedone for example by reversing the flow of heat transfer fluid so that itpasses from the cold tank, through boiler 470, and then to the hot tank,or by circulating the heat transfer fluid from the cold tank, throughboiler 470, and then back to the cold tank, or in any other suitablemanner.

For example, in some variations the heat transfer fluid (e.g., CaloriaHT-43) may be heated to about 300° C. by the solar field and used toproduce steam at about 35 bar in boiler 470 and/or stored in hot tank510. To continue power generation without coincident solar thermal inputfrom the solar field to the steam system (for example at night) the heattransfer fluid is circulated from the hot tank through boiler 470, wherethe heat transfer fluid may be cooled to about 250° C., for example, andthen stored in cold tank 530. The steam from boiler 470 and from MTHRU230 is mixed and superheated in HTHRU 210 from about 250° C. to about415° C., for example, and then expanded through steam turbine generator410 to produce power. For example, boiler 470 may produce about 126,000pounds/hour of saturated steam which is mixed with about 21,000pounds/hour of saturated steam produced by MTHRU 230, superheated, andthen expanded through the steam turbine generator to produce about 14.6MW of electric power.

When the volume of hot heat transfer fluid stored in hot tank 510 isconsumed, steam turbine generator 410 can continue to operate at lowerpower using the steam produced in MTHRU 230, optionally plus someadditional steam produced by a desuperheater used to limit the steamtemperature. As noted above, approximately the same volume of steamwould enter the steam turbine generator, so its throttle pressure wouldbe reduced to about 6 bar, for example, and the steam turbine generatorwould produce less power, for example about 2 MW.

Alternatively, or in addition, “cold” heat transfer fluid in cold tank530 (at about 250° C., for example) could be used to generate more steamat a lower pressure in boiler 470 in parallel with MTHRU 230 to increasethe steam turbine generator power output. For example, boiler 470 mayproduce about 2200 pounds/hour of saturated steam at about 11.4 bar,which would then be mixed with steam from MTHRU 230, also at about 11.4bar and superheated about 415° C. and expanded through the steam turbinegenerator to produce about 4 MW.

The flow rate of steam must satisfy the laws of thermodynamics, whichestablishes a pinch point, or minimum approach temperature differencebetween the heat transfer fluid and the steam in boiler 470 asillustrated in FIG. 4, which shows the temperature of steam and heattransfer fluid for the nameplate conditions of Tables 1A-1G and apart-load condition shown in Tables 3A-3G. In FIG. 4, plot 360 is for62.2 bar steam, plot 370 is for heat transfer fluid in a first passdischarge of heat from thermal energy storage (e.g., from the hot tank),plot 380 is for 32 bar steam, and plot 390 is for heat transfer fluid ina second pass discharge of heat from thermal energy storage (e.g., fromthe cold tank).

At part-load conditions, an additional mechanical constraint is imposed,namely that the volumetric flow rate into steam turbine generator 410 isapproximately unchanged from the nameplate condition, requiring bothlower steam pressure, density, and mass flow rate. With less steam massflow through heat recovery system 200, the gas temperatures are higher,and the steam flowing through HTHRU 210 may require de-superheating toavoid exceeding metal temperatures. Desuperheating, using high pressuresubcooled water from the boiler feedwater pump 460 is a common practice,and is not shown in FIGS. 2 and 3 for simplification.

The pinch point and flow rate of heat transfer fluid to boiler 470 fromthe hot tank (in the case of Tables 1A-1G) or cold tank (in the case ofTables 3A-3G) or cold tanks to the boiler 470 is the same. Theseconditions result in equal duration of storage and operating time at thefull load and part-load conditions. Alternatively, the part-load steampressure and steam turbine generator power output could be furtherreduced, which would allow the flow rate of heat transfer fluid to bereduced, thereby extending the duration of part-load operation stillfurther. Many combinations of part-load pressure and operating durationare possible, subject to the thermodynamic and volumetric flowconstraints.

There are also heat transfer constraints to be satisfied under the twooperating conditions, that is, there must be sufficient heat transferunder both the nameplate and part-load conditions. Newton's law ofcooling states that the heat transfer rate (kW) is promotional to thetemperature difference times the heat transfer area. The coefficient ofproportionality, or Heat Transfer coefficient, is largely dependent onReynolds Number, which is roughly constant because the volumetric flowrate is constant during sliding pressure operation. Accordingly, heattransfer performance at the part-load condition is not constraining.

FIG. 5 shows a schematic of an example configuration of thermal energystorage system 500, with three-way valves 550A-E inserted to direct theflow of heat transfer fluid to and from storage tanks 510 and 530. Thethree-way valves are illustrative, and alternative arrangements ofvalves and manifolds could be used to achieve the same result. Thediagram shows the “normal” flow arrangement previously described andshown in FIGS. 2 and 3, with cold heat transfer fluid flowing from coldtank 530 through valve 550D to pump 540 and thence to a solar field forheating. The heated heat transfer fluid returns from the solar field andflows through valve 550A into hot tank 510. When the solar field isunavailable, pump 540 would be turned off.

To produce steam at the nameplate pressure, heat transfer fluid wouldflow from hot tank 510 through valve 550C to pump 520, through boiler470. The cooled heat transfer fluid then returns to cold tank 530 viavalve 550E and valve 550B. To operate at part-load when the heattransfer fluid is depleted in hot tank 510, the valve positions would bechanged, so that heat transfer fluid would flow out of the cold tank530, through valve 550D to pump 520 and boiler 470, to return via valve550E and valve 550A to hot tank 550A.

It would be possible to continue heat storage discharge operation atstill lower saturated steam temperature and steam pressure by restoringthe valves to the original configuration.

Valve 550E allows heat transfer fluid to bypass boiler 470, formaintenance purposes, such as transferring heat transfer fluid betweentanks.

The arrangement described in FIG. 5 applies to Plant 10 or Plant 20 orany number of variants. Three or more tanks could be used in around-robin fashion, or multiple tanks could be operated in parallelarrangements. For example, a multiple tank system might dedicate thelowest temperature tank to feedwater heating, and reserve highertemperature tanks for boiling steam in order to improve the exergeticefficiency.

Moreover, this multiple pass approach can extend the operation ofthermal energy storage systems employing media other than low vaporpressure heat transfer fluids, including for example any of the examplethermal energy storage media described above. The temperatures at whichheat is discharged from the thermal energy system, the range oftemperatures over which discharge occurs, and the number of passes(i.e., the number of discharge cycles before the heat storage isrecharged) may depend on the particular heat storage medium used.

The heat rate for various operating conditions of one example of powerplant 20 is tabulated in Table 2 below. In this table “DSCC” indicatesoperation of the combustion turbine generator and also operation of thesteam turbine generator using heat from “hot” heat transfer fluid fromthe solar field and/or heat from thermal energy storage in combinationwith heat from the combustion turbine generator, “CTG only” indicatesthat the steam turbine generator is not operated, “CTG+HRSG” indicatesoperation of the combustion turbine generator and also operation of thesteam turbine generator using heat only from the combustion turbinegenerator and not from the solar field or thermal energy storage, and“Cold Tank Use” indicates operation of the combustion turbine generatorand also operation of the steam turbine generator using heat from “cold”heat transfer fluid from thermal energy storage in combination with heatfrom the combustion turbine generator. As shown in Table 2, the heatrate for power plant 20 may be exceptionally low in normal operation.Operating at part load as described above (“Cold Tank Use” or“CTG+HRSG”), the heat rate may be comparable to that of a modernconventional combined cycle power plant. Accordingly, the arrangementand operation of DSCC power plant 20 described above can increase theplant capacity and the utilization of thermal energy storage and improvethe heat rate.

As noted above, although plant 10 and plant 20 employ a solar thermalsystem 300 as a heat source for their bottoming cycles, any othersuitable heat source may be used instead of solar thermal energy.Different heat sources may store heat in the thermal energy storagesystem at different maximum storage temperatures. Combinations of heatsources, thermal energy storage system heat storage media, and thermalenergy storage system configuration may be selected to produce and storeheat at desired temperatures. The bottoming cycle may then be operatedat a sequence of temperatures and pressures to advantageously use thestored heat, similarly to plant 10 and plant 20 as described above.

As an additional example DSCC power plant, FIG. 6 shows a power plant 30that is very similar to power plant 20 of FIG. 3 in configuration andoperation, except that power plant 30 comprises an electric heatingsystem 350 in place of solar thermal system 300.

Electric heating system 350 uses resistance or inductive heating totransform electric energy to thermal energy for storage in hightemperature energy storage system 500. Electricity may be delivered toelectric heating system 350 from any suitable power generation sourcesuch as, for example, a fossil fuel fired power plant, a nuclear powerplant, a wind turbine power plant, and a thermal or solar photovoltaicsolar power plant. The electricity may be produced adjacent to plant 30,at a distance from plant 30 and transmitted to plant 30 over anelectrical grid, or may be a combination of local and remotely generatedpower. Electric power may be produced or transmitted at any suitablevoltage as either alternating current (AC) or direct current (DC).Electric heating system 350 may use the power in the form delivered, ormay transform or convert it to a voltage and waveform more suitable forthe control of power delivery.

The flow of heat transfer fluid from cold tank 530 to electric heatingsystem 350 may be regulated in proportion to the electric powerdelivered to electric heating system 350 by pump 540 in order toestablish a desired temperature of the heat transfer fluid returning tohot tank 510. The regulation of flow rate may be performed by varyingthe speed of pump 540, or by the use of control valves to recirculateexcess heat transfer fluid back to cold tank 530. With the methods justdescribed, the electric heating system 350 may readily storetime-varying quantities of electric energy to facilitate balancing ofelectric supply and demand.

The temperature of the heat transfer fluid returning to hot tank 510 mayalso be regulated by adjusting the amount of electric energy deliveredto electric heating system 350, using any suitable temperature controlmeans, such as connecting or interrupting electric heaters (also knownas on-off control), or proportional control means such as varying thevoltage and/or current delivered to electric resistance heaters withinelectric heating system 350.

Electric heating system 350 may be used to achieve higher temperaturesthan would be possible with a solar thermal system 300 based onoil-based parabolic trough technology, which is typically limited toabout 390° C. Electric heating system 350 may, for example, heat moltensalt to a temperature of about 565° C., as practiced with some tower andlinear solar thermal systems. Electric heating system 350 could also beused to store energy at very high temperature, for example in the phasechange of metals, as is done in electric arc furnaces for steel making,which achieve temperature of 1800° C. or more.

The storage temperature is constrained by the materials limitations ofboth the storage medium and the container and supporting systems. Forexample, a molten salt may be chemically unstable at high temperatures,which limits solar thermal storage to about 565° C. At high temperature,molten salts may also be corrosive to steels used in tanks, piping,pumps, heaters, valves, and instruments, which might necessitate the useof expensive alternate materials. Accordingly, reliability and economicconsiderations may limit the temperature of the heat transfer fluidexiting electric heating system 350.

Molten salts used as heat transfer fluids are also constrained at lowertemperatures where viscosity increases pumping power requirements andthere is a risk of freezing. Molten salt formulations used in towerconcentrating solar power applications have a typical operatingtemperature range of 250° C. to 565° C.; medium temperature formulationsare available with an operating temperature range of 140° C. to 485° C.Either formulation would permit a higher efficiency DSCC implementationat 70 bar rated steam pressure, with the simple non-reheatconfiguration. The 250° C. operating limit of the higher temperaturesalt would limit sliding pressure operation to about 40 bar, whereas thelower temperature salt would permit operation down to about 4 bar.

The higher temperature produced by electric heating system 350 increasesthe density of energy storage within hot tank 510, which may permit asmaller and less costly tank, and a smaller quantity of storage media. Afigure of merit in evaluating heat transfer fluids for energy storage isthe product of density and specific heat, which has units of kiloJoulesper cubic meter per degree. Accordingly, increasing the temperaturerange over which the storage operates also increases the useful storedenergy. Parameters of some commercially available heat transfer fluidsand heat storage media are shown below in Table 4.

In addition to using a heat transfer fluid as a heat storage medium, asnoted above heat transfer fluids can be used to transfer heat into orout of solid heat storage media, such as the concrete media availablefrom EnergyNest or the graphite media available from Graphite Energy.The figure of merit of the EnergyNest system in Table 4 was derived fromthe energy claimed to be stored within their module over the entiretemperature range. To take full advantage of the high temperaturecapacity of graphite, the graphite heat storage modules could beelectrically heated, for example by embedding resistive heatingelements. The additional heat stored at high temperature could betransferred to boiler 470 by reducing the flow rate of heat transferfluid from hot tank 510 and employing a higher temperature drop beforereturning heat transfer fluid to cold tank 530. Alternatively, the flowrate of heat transfer fluid could be maintained approximately constantduring both the normal operating pressure case and the part load case,as shown in Table 1A and Table 3A.

More generally, a method of operating a DSCC power plant such as powerplant 10, power plant 20, or power plant 30, for example, may compriseoperating a combustion turbine generator to generate electricity andproduce hot exhaust gases, and storing heat from a heat source otherthan the combustion turbine generator in a thermal energy storage systemat a temperature T1. In a first mode of operation the method maycomprise producing steam in a first boiler by heating feedwater withheat supplied at temperature T1 from the thermal energy storage systemby a heat transfer fluid, thereby cooling the heat transfer fluid,storing heat from the cooled heat transfer fluid in the thermal energystorage system at a temperature T2<T1, heating the steam from the firstboiler with heat from the combustion turbine exhaust gases to producesuperheated steam at pressure P1, and expanding the superheated steam atpressure P1 through the steam turbine generator to generate electricity.

In a second mode of operation, after depleting the thermal energystorage system of heat stored at temperature T1, the method may compriseproducing steam in the first boiler by heating feedwater with heatsupplied from the thermal energy storage system at a temperature of T2or less by a heat transfer fluid, heating the steam from the firstboiler with heat from the combustion turbine exhaust gases to producesuperheated steam at pressure P2<P1, and expanding the superheated steamat pressure P2 through the steam turbine generator to generateelectricity. The temperature of the superheated steam may be controlledor limited by suitable means, such as a water spray attemperator.

The pressures P1 and P2 may be controlled by the rate at which heat issupplied from the thermal energy storage system to the boiler, forexample by controlling a heat transfer fluid flow rate to control therate at which heat is supplied from the thermal energy storage system tothe boiler.

In both the first mode of operation and the second mode of operation,the method may comprise preheating the feedwater with heat from thecombustion turbine exhaust gases.

The first mode of operation may comprise producing steam in a secondboiler operated in parallel with the first boiler by heating feedwaterwith heat from the combustion turbine exhaust gases, mixing the steamfrom the first boiler with the steam from the second boiler, and heatingthe mixture of steam from the first boiler and steam from the secondboiler with heat from the combustion turbine exhaust gases to producethe superheated steam at pressure P1. Similarly, the second mode ofoperation may comprise producing steam in the second boiler operated inparallel with the first boiler by heating feedwater with heat from thecombustion turbine exhaust gases, mixing the steam from the first boilerwith the steam from the second boiler, and heating the mixture of steamfrom the first boiler and steam from the second boiler with heat fromthe combustion turbine exhaust gases to produce the superheated steam atpressure P2. Optionally, in these variations of the first and secondmodes of operation the first boiler may produce steam using heatexclusively from the thermal energy storage system and the second boilermay produce steam using heat exclusively from the combustion turbineexhaust gases.

The method may comprise a third mode of operation comprising supplyingheat from the heat source to the first boiler at the temperature T1,without first storing the heat in the thermal energy storage system, toproduce steam by heating feedwater with the heat supplied at temperatureT1, heating the steam from the first boiler with heat from thecombustion turbine exhaust gases to produce superheated steam atpressure P1, and expanding the superheated steam at pressure P1 throughthe steam turbine generator to generate electricity.

In the third mode of operation the method may comprise preheating thefeedwater with heat from the combustion turbine exhaust gases.

The third mode of operation may comprise producing steam in a secondboiler operated in parallel with the first boiler by heating feedwaterwith heat from the combustion turbine exhaust gases, mixing the steamfrom the first boiler with the steam from the second boiler, and heatingthe mixture of steam from the first boiler and steam from the secondboiler with heat from the combustion turbine exhaust gases to producethe superheated steam at pressure P1. Optionally, in this variation ofthe third mode of operation the first boiler may produce steam usingheat exclusively from the thermal energy storage system and the secondboiler may produce steam using heat exclusively from the combustionturbine exhaust gases.

The method may comprise a fourth mode of operation comprising producingsteam by heating feedwater with heat from the combustion turbine exhaustgases and without using heat from the heat source or the thermal energystorage system, heating the steam with heat from the combustion turbineexhaust gases and without using heat from the heat source or the thermalenergy storage system to produce superheated steam at pressure P3<P1,and expanding only the superheated steam at pressure P3, no othersuperheated steam, through the steam turbine generator to generateelectricity.

The method may comprise starting operation of the first boiler and thesteam turbine generator using heat from the heat source, the thermalenergy storage system, or the heat source and the thermal energy storagesystem before starting operation of the combustion turbine generator.This could reduce startup fuel consumption and air emissions.

The heat source may collect solar energy as heat. The heat source maycomprise an electric heater. In the latter case, the method may comprisepowering the electric heater with electricity generated withphotovoltaic solar cells, with a solar thermal electric power plant,with a wind turbine, with a hydroelectric power plant, using nuclearpower, by a fossil fuel fired thermal power plant, by a geothermal powerplant, by a combination thereof, or with electricity generated in anyother suitable manner.

Storing the heat in the thermal energy storage system may comprisestoring the heat in a low vapor pressure heat transfer fluid at apressure of about one atmosphere. In such a case, the parameters T1, T2,P1, and P2 may have, for example, the following values: 300° C.≤T1≤340°C., 35 bar≤P1≤75 bar, 240° C.≤T2≤290° C., and 15 bar≤P2≤55 bar.

Storing the heat in the thermal energy storage system may comprisestoring the heat in a molten salt. In such a case, the parameters T1,T2, P1, and P2 may have, for example, the following values: 350°C.≤T1≤600° C., 65 bar≤P1≤125 bar, 250° C.≤T2≤400° C., and 30 bar≤P2≤90bar.

Storing the heat in the thermal energy storage system may comprisestoring the heat in a solid heat storage medium. In such a case, theparameters T1, T2, P1, and P2 may have, for example, the followingvalues: 350° C.≤T1≤600° C., 65 bar≤P1≤125 bar, 250° C.≤T2≤400° C., and30 bar≤P2≤90 bar.

More generally, the parameters T1, T2, P1, and P2 may have, for example,the following values: 300° C.≤T1≤600° C., 35 bar≤P1≤125 bar, 240°C.≤T2≤400° C., and 15 bar≤P2≤90 bar.

For a DSCC hybrid solar thermal combined cycle power plant with eighthours of thermal energy storage, the power plant could run around theclock during summer periods to achieve 88% Capacity Factor:

-   -   8 full-power hours while the sun shines, and heat transfer fluid        flows from the solar field to the Steam Generator and to the hot        tank for storage;    -   8 full-power hours with heat transfer fluid flowing from the hot        tank to boiler 470 and then to the cold tank; and    -   8 part-power hours with heat transfer fluid flowing from the        cold tank to the boiler 470 and then to the hot tank, or back to        the cold tank.

Likewise, in winter periods, the operating hours would be extended.

Various tank and piping arrangements may be used to increase capacityfactor and/or reduce cost as described above. The moderate temperatureof the heat transfer fluid in those DSCC systems described herein thatemploy low vapor pressure heat transfer fluids may allow either tank tocontain hot heat transfer fluid. Accordingly the hot and cold storagefunctions can alternate between two tanks, or three or more tanks couldbe used in a round-robin fashion, with one tank receiving heat transferfluid and the others containing heat transfer fluid at highertemperatures.

Optimal operation of solar hybrid power plant 10 and 20 depends onperformance parameters (for example, storage capacity, solar fieldcapacity, combustion turbine generator capacity) which must beestablished during design, as well as parameters that are variableduring the operation of the power plant. In addition, the solar resourceforecast, instantaneous resource quantity, the available storagecapacity, and the power capacity commitment and/or merchant power priceand fuel cost are factors affecting the optimal use of the overall powerplant. Moreover, assumptions about these latter factors may influencethe selection of the performance design parameters.

For explanation, consider how certain parameters and factors wouldaffect the optimal operation of the power plant operating on a gridsubject to the ‘Duck Curve.’ Suppose the configuration of the powerplant described above with operating parameters listed in Table 1, withfour full-power hours of thermal storage (1700 MW-hour). On a particularday, eight full power hours of solar thermal resource (4250 MW-hour) maybe forecast. During the mid-day period, when the highest solar resourceis available, and the ‘Duck Curve’ overgeneration risk is highest, itmay be desirable to reduce generation and divert solar thermal energy tostorage. But around solar noon, the solar thermal system might have acapacity of 600 MW, meaning that the storage would be fully re-chargedwithin 3 hours.

In anticipation of the need for storing energy in the thermal energystorage system, the optimal operating plan may attempt to fully depletethe energy storage during the early morning, and then split hot heattransfer fluid between the boiler and storage, with the intention thatstorage becomes full just as the solar resource could no longer supplythe full boiler capacity. Given that the solar resource is differenteach day of the year, this is a challenge to achieve, once a fixedamount of storage has been selected. Accordingly, the design parametersmay be optimized to maximize the value of the storage and solar thermalsystems over the life of the power plant. Once a design point has beenselected, the operating strategy may then be optimized to achieve themaximum revenue, given the fuel and electricity tariffs, which may varyby time of day.

Combustion turbine generators have the capability of burning a varietyof fuels in addition to natural gas. In the event an emergencyinterrupts the supply of fuel gas, heat transfer fluid could be burnedin the combustion turbine generator to maintain power generation.Although such an unexpected event (triggered for example by a naturaldisaster such as an earthquake) might never occur, the solar powerplants described in this specification may readily facilitate emergencyoperation, upon provision of suitable fuel oil piping, nozzles,atomizers, etc.

This disclosure is illustrative and not limiting. Further modificationswill be apparent to one skilled in the art in light of this disclosureand are intended to fall within the scope of the appended claims. Forexample, heat from a thermal energy storage system may be used toproduce steam to maintain condenser seals and/or to peg a deaeratingfeedwater heater.

TABLE 1A Gas flows through CTG 100 and HRSG 200 Flow (kg/s) PressureTemperature Other (bar) (° C.) Total O2 N2 H2O CH4 CO2 gases Fuel to CTG100 28.000 335.477 14.0 0.0 0.3719 0.0 12.7 0.26 balance Inlet air toCTG 100 1.000 15.0 590.0 136.5 445.6 0.0 0.0 0.3 balance CTG 100 Exhaust1.002 615.1 604.0 83.4 445.9 29.7 0.0 37.4 balance HTHRU 210 Exhaust1.001 346.7 604.0 83.4 445.9 29.7 0.0 37.4 balance MTHRU 230 Exhaust1.001 169.0 604.0 83.4 445.9 29.7 0.0 37.4 balance LTHRU 240 Exhaust1.000 66.5 604.0 83.4 445.9 29.7 0.0 37.4 balance

TABLE 1B Steam and Water Flows Flow Pressure Temperature rate Enthalpy(bar) (° C.) (kg/s) (kJ/kg) Superheated Steam from 61.5 548.0 250 3535.0HTHRU 210 Exhaust From STG 410 0.085 42.7 250 2534.6 Condensate from ACC420 0.085 40.0 250 167.5 Condensate from Transfer 2.21 40.1 250 168.2Pump 430 Condensate from Deaerator 2.21 104.9 250 439.6 450 Feedwaterfrom Pump 460 62.2 105.8 250 448.0 Steam from MTHRU 230 62.2 280.0 502792.1 Steam From Boiler 470 62.2 280.0 200 2792.1

TABLE 1C HTF flows between Storage System 500 and Boiler 470 PressureTemperature Flow rate (bar) (° C.) (kg/s) HTF from Hot Tank 510 1 3322358 HTF from Pump 520 to boiler 470 10 332 2358 HTF from boiler 470 toCold 3 261 2358 Tank 530

TABLE 1D HTF flows between LTHRU 240 and Deaerator 450 PressureTemperature Flow rate (bar) (° C.) (kg/s) HTF from LTHRU 240 5 140 581HTF from Pump 520 to boiler 470 10 140 581 HTF from boiler 470 to Cold 390 581 Tank 530

TABLE 1E Air Flow through Air Cooled Condenser 420 Pressure TemperatureFlow rate Enthalpy Stream (bar) (° C.) (kg/s) (kJ/kg) Cooling Air IN1.00 15.0 30000.00 15.30 Cold Air to ACC 1.00 15.0 30000.00 15.33 HotAir from ACC 1.00 34.5 30000.00 35.05

TABLE 1F Heating Duties Power Heat Release Rate in Combustor of CTG 100(HHV) 739.5 HTHRU 210 185.1 Boiler 470 469.0 MTHRU 230 117.2 LTHRU 24067.9 ACC 420 591.8 Deaerator 450 67.9

TABLE 1G Electric Generation and Efficiency Power Component CTG 100Power Output 273.5 STG 410 Power Output 250.1 Gross Power Output (CTG100 plus STG 410) 523.6 Fuel Consumption 739.5 Gross Efficiency 70.8%Principal Parasitic Loads Fuel Gas Compressor (not shown) 11.7 AirCooled Condenser 420 1.0 Boiler Feedwater Pump 460 2.0 HTF CirculatingPump 520 4.5 HTF Circulating Pump 610 0.5 Other House Loads 1 NetGeneration 502.9 Net Efficiency 68.0%

TABLE 2 Power Plant Performance Gross Gross HHV Fuel Heat Case Power(kW) Fuel (MMBtu/h) Rate (Btu/kWh) DSCC 29600 138 4662 CTG only 15000138 9200 CTG + HRSG 17000 138 8117 Cold Tank Use 19000 138 7263

TABLE 3A Gas flows through CTG 100 and HRSG 200 Flow (kg/s) PressureTemperature Other (bar) (° C.) Total O2 N2 H2O CH4 CO2 gases Fuel to CTG100 28.000 335.477 14.0 0.0 0.3719 0.0 12.7 0.26 balance Inlet air toCTG 100 1.000 15.0 590.0 136.5 445.6 0.0 0.0 0.3 balance CTG 100 Exhaust1.002 615.1 604.0 83.4 445.9 29.7 0.0 37.4 balance HTHRU 210 Exhaust1.001 460.1 604.0 83.4 445.9 29.7 0.0 37.4 balance MTHRU 230 Exhaust1.001 284.90 604.0 83.4 445.9 29.7 0.0 37.4 balance LTHRU 240 Exhaust1.000 233.4 604.0 83.4 445.9 29.7 0.0 37.4 balance

TABLE 3B Steam and Water Flows Flow Pressure Temperature rate Enthalpy(bar) (° C.) (kg/s) (kJ/kg) Superheated Steam from 31.3 546.0 125 3559.6HTHRU 210 Exhaust From STG 410 0.085 42.7 125 2534.6 Condensate from ACC420 0.085 40.0 125 167.5 Condensate from Transfer 2.21 40.1 125 168.2Pump 430 Condensate from Deaerator 2.21 104.9 125 439.6 450 Feedwaterfrom Pump 460 32.0 105.8 125 443.7 Steam from MTHRU 230 32.0 238.0 502806.0 Steam From Boiler 470 32.0 238.0 69.1 2806.0 Desuperheating Sprayto 32.0 105.8 5.9 443.7 HTHRU 210

TABLE 3C HTF flows between Storage System 500 and Boiler 470 PressureTemperature Flow rate (bar) (° C.) (kg/s) HTF from Cold Tank 530 1 2612358 HTF from Pump 520 to boiler 470 10 261 2358 HTF from boiler 470 toHot 3 238.5 2358 Tank 510

TABLE 3D HTF flows between LTHRU 240 and Deaerator 450 PressureTemperature Flow rate (bar) (° C.) (kg/s) HTF from LTHRU 240 5 140 290HTF from Pump 520 to boiler 470 10 140 290 HTF from boiler 470 to Cold 390 290 Tank 530

TABLE 3E Air Flow through Air Cooled Condenser 420 Pressure TemperatureFlow rate Enthalpy Stream (bar) (° C.) (kg/s) (kJ/kg) Cooling Air IN1.00 15.0 15000.00 15.30 Cold Air to ACC 1.00 15.0 15000.00 15.33 HotAir from ACC 1.00 34.5 15000.00 35.05

TABLE 3F Heating Duties Power Heat Release Rate in Combustor of CTG 100(HHV) 739.5 HTHRU 210 108.2 Boiler 470 163.1 MTHRU 230 118.1 LTHRU 24033.9 ACC 420 306.5 Deaerator 450 33.9

TABLE 3G Electric Generation and Efficiency Power Component CTG 100Power Output 273.5 STG 410 Power Output 117.5 Gross Power Output (CTG100 plus STG 410) 391.0 Fuel Consumption 739.5 Gross Efficiency 52.9%Principal Parasitic Loads Fuel Gas Compressor (not shown) 11.7 AirCooled Condenser 420 .5 Boiler Feedwater Pump 460 1.0 HTF CirculatingPump 520 4.5 HTF Circulating Pump 610 0.5 Other House Loads 1 NetGeneration 371.8 Net Efficiency 50.3%

TABLE 4 Heat Transfer Fluids and Heat Storage Media Temperature StorageMedium Range Density Specific Heat Figure of Merit Duratherm 630  −5C.-335 C. 651.46 kg/cu.m.   2.971 kJ/kg-C.  1935 kJ/cu.m.-C. DynaleneMS-2 140 C.-485 C. 1890 kg/cu.m  1.59 kJ/kg-C. 3005 kJ/cu.m.-C. DynaleneMS-1 250 C.-565 C. 1900 kg/cu.m. 1.40 kJ/kg-C. 2660 kJ/cu.m.-C.EnergyNest  0 C.-427 C.  353 kJ/cu.m.-C. Graphite   0 C.-2000 C. 2090kg/cu.m. 0.71 kJ/kg-C. 1484 kJ/cu.m.-C.

What is claimed is:
 1. A method of operating a combined cycle electricpower plant, the method comprising: operating a combustion turbinegenerator to generate electricity and produce hot exhaust gases; storingheat from a heat source other than the combustion turbine generator in athermal energy storage system at a temperature T1; in a first mode ofoperation: producing steam in a first boiler by heating feedwater withheat supplied at temperature T1 from the thermal energy storage systemby a heat transfer fluid, thereby cooling the heat transfer fluid, andstoring heat from the cooled heat transfer fluid in the thermal energystorage system at a temperature T2<T1; heating the steam from the firstboiler with heat from the combustion turbine exhaust gases to producesuperheated steam at pressure P1; and expanding the superheated steam atpressure P1 through the steam turbine generator to generate electricity;and in a second mode of operation, after depleting the thermal energystorage system of heat stored at temperature T1: producing steam in thefirst boiler by heating feedwater with heat supplied from the thermalenergy storage system at a temperature of T2 or less by a heat transferfluid; heating the steam from the first boiler with heat from thecombustion turbine exhaust gases to produce superheated steam atpressure P2<P1; and expanding the superheated steam at pressure P2through the steam turbine generator to generate electricity.
 2. Themethod of claim 1, wherein the pressure P1 and the pressure P2 are eachcontrolled by a rate at which heat is supplied from the thermal energystorage system to the boiler.
 3. The method of claim 2, comprisingcontrolling a heat transfer fluid flow rate to control the rate at whichheat is supplied from the thermal energy storage system to the boiler.4. The method of claim 1 comprising, in both the first mode of operationand the second mode of operation, preheating the feedwater with heatfrom the combustion turbine exhaust gases.
 5. The method of claim 1comprising: in the first mode of operation: producing steam in a secondboiler operated in parallel with the first boiler by heating feedwaterwith heat from the combustion turbine exhaust gases; mixing the steamfrom the first boiler with the steam from the second boiler; and heatingthe mixture of steam from the first boiler and steam from the secondboiler with heat from the combustion turbine exhaust gases to producethe superheated steam at pressure P1. and in the second mode ofoperation: producing steam in the second boiler operated in parallelwith the first boiler by heating feedwater with heat from the combustionturbine exhaust gases; mixing the steam from the first boiler with thesteam from the second boiler; and heating the mixture of steam from thefirst boiler and steam from the second boiler with heat from thecombustion turbine exhaust gases to produce the superheated steam atpressure P2.
 6. The method of claim 5, wherein in the first and secondmodes of operation the first boiler produces steam using heatexclusively from the thermal energy storage system and the second boilerproduces steam using heat exclusively from the combustion turbineexhaust gases.
 7. The method of claim 1, comprising: in a third mode ofoperation: supplying heat from the heat source to the first boiler atthe temperature T1, without first storing the heat in the thermal energystorage system, to produce steam by heating feedwater with the heatsupplied at temperature T1; heating the steam from the first boiler withheat from the combustion turbine exhaust gases to produce superheatedsteam at pressure P1; and expanding the superheated steam at pressure P1through the steam turbine generator to generate electricity.
 8. Themethod of claim 7, comprising in the third mode of operation preheatingthe feedwater with heat from the combustion turbine exhaust gases. 9.The method of claim 7, comprising in the third mode of operation:producing steam in a second boiler operated in parallel with the firstboiler by heating feedwater with heat from the combustion turbineexhaust gases; mixing the steam from the first boiler with the steamfrom the second boiler; and heating the mixture of steam from the firstboiler and steam from the second boiler with heat from the combustionturbine exhaust gases to produce the superheated steam at pressure P1.10. The method of claim 9, wherein in the third mode of operation thefirst boiler produces steam using heat exclusively from the heat sourceand the second boiler produces steam using heat exclusively from thecombustion turbine exhaust gases.
 11. The method of claim 1, comprising:in a fourth mode of operation: producing steam by heating feedwater withheat from the combustion turbine exhaust gases and without using heatfrom the heat source or the thermal energy storage system; heating thesteam with heat from the combustion turbine exhaust gases and withoutusing heat from the heat source or the thermal energy storage system toproduce superheated steam at pressure P3<P1; and expanding only thesuperheated steam at pressure P3, no other superheated steam, throughthe steam turbine generator to generate electricity.
 12. The method ofclaim 1, wherein the heat source collects solar energy as heat.
 13. Themethod of claim 1, wherein the heat source comprises an electric heater.14. The method of claim 13, comprising powering the electric heater withelectricity generated with photovoltaic solar cells, with a solarthermal electric power plant, with a wind turbine, with a hydroelectricpower plant, using nuclear power, by combusting a fossil fuel, or by acombination thereof.
 15. The method of claim 1, wherein storing the heatin the thermal energy storage system comprises storing the heat in a lowvapor pressure heat transfer fluid at a pressure of about oneatmosphere.
 16. The method of claim 15, wherein: 300° C.≤T1≤340° C.; 35bar≤P1≤75 bar; 240° C.≤T2≤290° C.; and 15 bar≤P2≤55 bar.
 17. The methodof claim 1, wherein storing the heat in the thermal energy storagesystem comprises storing the heat in a molten salt.
 18. The method ofclaim 17, wherein: 350° C.≤T1≤600° C.; 65 bar≤P1≤125 bar; 250°C.≤T2≤400° C.; and 30 bar≤P2≤90 bar.
 19. The method of claim 1, whereinstoring the heat in the thermal energy storage system comprises storingthe heat in a solid heat storage medium.
 20. The method of claim 19,wherein: 350° C.≤T1≤600° C.; 65 bar≤P1≤125 bar; 250° C.≤T2≤400° C.; and30 bar≤P2≤90 bar.
 21. The method of claim 1, wherein: 300° C.≤T1≤600°C.; 35 bar≤P1≤125 bar; 240° C.≤T2≤400° C.; and 15 bar≤P2≤90 bar.
 22. Themethod of claim 1, comprising starting operation of the first boiler andthe steam turbine generator using heat from the heat source, the thermalenergy storage system, or the heat source and the thermal energy storagesystem before starting operation of the combustion turbine generator.